JP2017512277A - Gaseous fuel combustion system for internal combustion engines - Google Patents

Abstract

Diesel cycle engines are known which have greater power, torque and efficiency than Otto cycle engines of similar displacement. When the fuel is a gaseous fuel such as natural gas, a pilot fuel (such as diesel) is usually required to facilitate ignition in a gas fueled diesel cycle engine. It would be advantageous to reduce the output, torque and efficiency gap between a diesel cycle engine and a gas fueled Otto cycle engine. A combustion apparatus for a gas fuel type internal combustion engine is provided with a combustion chamber defined by a cylinder bore, a cylinder head, and a piston that reciprocates within the cylinder bore. The diameter of the cylinder bore is at least 90 mm and the ratio between this diameter and the stroke length of the piston is at most 0.95. At least one intake flow path for supplying charge to the combustion chamber is provided, and at least one intake valve is formed in the cylinder head and cooperates with the intake flow path to produce a significant tumble flow motion in the combustion chamber. To do. [Selection] Figure 2

Description

  The present invention relates to a combustion apparatus for a gas fuel type internal combustion engine (internal combustion engine).

  The flow of intake charge has a great influence on the performance of a gaseous fuel internal combustion engine. Mixing air, possibly exhaust gas, and gaseous fuel affects the combustion characteristics in the combustion chamber. The charge movement in the combustion chamber during the intake stroke (intake stroke) and during the subsequent compression stroke determines the mixing level and characteristics of the gaseous fuel. In some parts of the engine map, it is desirable to have a uniform air-fuel charge, while in other parts of the engine map the engine charge is improved by stratifying the fuel charge near the igniter. In yet another part of the engine map, the engine performance is further improved by locally enriching the air-fuel mixture and making it globally thinner. The generation of intense turbulence is an important factor for stabilizing the ignition process and for increasing the propagation speed (flame speed) in front of the flame, particularly in the case of lean combustion. Two techniques for generating charge motion in a cylinder are known as tumble motion and swirl motion. Tumble motion and swirl motion can be represented by dimensionless numbers used to quantify rotational and angular motion inside the cylinder, known as tumble ratio and swirl ratio, respectively. Both of these values are calculated as the effective angular velocity of the air motion in the cylinder divided by the engine speed.

  It is known to use tumble motion for a direct injection light duty (small) gasoline engine that employs fuel stratification surrounding the igniter. In the tumble motion, also called vertical swirl or barrel swirl, the rotation axis of the intake charge in the cylinder is orthogonal to the cylinder axis. In this specification, the light duty engine means an engine having a cylinder bore diameter (inner diameter) smaller than 90 millimeters (mm). Fuel stratification is an effective technique for extending the limits of lean combustion in spark ignition engines, thus increasing fuel savings and reducing emissions compared to conventional light duty gasoline engines. The tumble motion can be effective to increase the flow velocity in the vicinity of the wall surface, even relatively lean in the compression stroke, thereby facilitating evaporation of the fuel wall film formed by the fuel spray collision.

  U.S. Pat. No. 5,553,580 issued Sep. 10, 1996 (inventor: Ganoung) describes a gasoline engine employed to reduce the net fuel consumption for a light duty engine. A barrel stratified combustion chamber in a high squish area is disclosed. Two intake valves are fluidly connected to respective intake passages configured as tumble ports. By introducing gasoline into one of these intake channels, a barrel stratified charge is generated in the cylinder, and the stratified barrel swirl is located near an asymmetrically located spark plug (spark plug). It is formed. Barrel swirls do not increase the combustion rate, but rather help stratify the air-fuel charge in the cylinder when the spark plug is ignited. Large squish areas increase combustion speed by increasing turbulence intensity during combustion.

  It is known to use swirl motion for diesel cycle (compression ignition) heavy duty (large) engines. In the swirl motion, the rotation axis of the intake charge in the cylinder is the cylinder axis. As used herein, a heavy duty engine means an engine having a cylinder bore diameter greater than 120 millimeters (mm). Swirl motion is known to reduce particulate matter (PM) emissions from the engine. Compression ignition engines employ high injection pressures, which improve droplet breakup for liquid fuels, and improve air / fuel mixing during spraying with high injection pressures for both liquid and gaseous fuels In addition, there is a tendency to increase the turbulence intensity in the combustion chamber. This is especially important during transient conditions when the combustion system needs to deal with low air-fuel ratio conditions without releasing high PM. Employing swirl motion can reduce the PM generation action under certain transient conditions even when high injection pressure is used. In order to convert an engine designed for swirl motion into tumble motion, it is necessary to change the direction with respect to the intake flow path, which requires different cylinder heads. The need for a new cylinder head interferes with experiments using this technology for medium duty engines and large engines. The reason is that the diesel engine is already regarded as the most effective internal combustion engine.

  One goal of engine design is to reduce cylinder displacement with little loss in performance (horsepower and torque). As fuel costs and street congestion increase, vehicle drivers require smaller vehicles that are the same as large vehicles but provide the overall performance with fuel savings. Alternative gaseous fuels are increasingly finding new areas in the automotive market segment dominated by gasoline and diesel fuel engines in many jurisdictions. In the light duty field, port-injected natural gas engines have a long history in the aftermarket segment, and recently OEM versions of these vehicles have been introduced. In the heavy duty field, high pressure direct injection (HPDI) engine systems are adapted to the performance of diesel fuel engines and improve fuel efficiency compared to port injection natural gas engines.

US Pat. No. 5,553,580 US Provisional Patent Application No. 61 / 870,203

  There is a need for a gaseous fuel engine that has performance comparable to a large engine, but that has improved fuel economy for an engine designed especially for at least medium duty service.

  The improved combustion apparatus of the present invention for a gas-fueled internal combustion engine has a combustion chamber defined by a cylinder bore, a cylinder head, and a piston that reciprocates within the cylinder bore. The diameter of the cylinder bore is at least 90 mm and the ratio between this diameter and the stroke length of the piston is at most 0.95. Also, at least one intake passage for supplying charge to the combustion chamber is provided, and at least one intake valve is formed in the cylinder head and cooperates with the intake passage to cause significant tumble flow motion in the combustion chamber. Give birth.

  In a preferred embodiment of the present invention, the ratio is at least 0.75, the diameter is 120 mm or less, or both. The stroke volume of the cylinder bore is preferably in the range of 0.8 liters to 2.5 liters. An injection valve can be formed that introduces gaseous fuel upstream from at least one intake valve. Alternatively, the injection valve can be arranged in the combustion chamber so that gaseous fuel can be introduced directly into the combustion chamber. An ignition device may be placed in the combustion chamber to facilitate ignition of gaseous fuel and charge. In a preferred embodiment of the present invention, the ignition device is a spark plug. The tumble flow motion preferably has an average tumble ratio in the range of 2-5. The maximum engine speed of the internal combustion engine is 2700 rpm. Each intake valve has a valve member and a valve seat. The valve seat has a valve seat angle in the range of 25 ° to 35 °. In a preferred embodiment of the present invention, the valve seat angle is substantially 30 °. The difference between the valve seat angle and the port angle is in the range of -5 ° to 5 °. The compression ratio of the internal combustion engine is at least 11 to 1, and in the preferred embodiment is at most 15 to 1. The intake manifold includes a first distribution chamber in fluid communication with the intake air of the internal combustion engine, a second distribution chamber in fluid communication with at least one intake flow path, a first distribution chamber, and a second distribution chamber. And a diffuser for fluidly connecting the distribution chamber. The combustion apparatus may have an EGR valve that selectively supplies exhaust gas to the intake manifold. In a preferred embodiment of the present invention, the exhaust gas is cooled before being supplied to the intake manifold. A throttle valve can be employed to variably supply air to the intake manifold. In a preferred embodiment of the invention, the throttle valve is commanded to maintain a stoichiometric gaseous fuel-air mixture within a predetermined tolerance.

  In a preferred embodiment of the present invention, the at least one intake passage is a first intake passage and a second intake passage, and the at least one intake valve is a first intake valve and a second intake valve. . The combustion apparatus further includes an injection valve that receives the gaseous fuel, and a shunt that is in fluid communication with the first intake passage and the second intake passage that discharges the received gaseous fuel from the injection valve. . The shunt has a body having a bore and a pair of conduits. The bore is in fluid communication with the injection valve, and each conduit is in fluid communication with the bore and in fluid communication with one each of the first intake flow path and the second intake flow path.

  The internal combustion engine has an engine block and an intake manifold, and the at least one intake passage is a first intake passage and a second intake passage. In another preferred embodiment of the invention, the combustion device further comprises six bolts arranged around the cylinder bore, which hold the cylinder head against the engine block. The first intake flow path and the second intake flow path extend along both sides of one of the bolts from the intake manifold toward the combustion chamber.

  A new gaseous fuel shunt for diverting the flow of gaseous fuel from the fuel injector includes a body portion having a bore in fluid communication with the fuel injector, a first conduit and a second fluid in fluid communication with the bore. A conduit. The gaseous fuel flow is split into a first flow and a second flow in the first conduit and the second conduit, respectively. In a preferred embodiment of the invention, the fuel injection valve is part of the fuel injector and the bore is configured to accommodate the nozzle of the fuel injector. In other preferred embodiments, the first conduit and the second conduit are substantially perpendicular to the longitudinal axis of the bore, or the body and the first conduit and the second conduit are integrated components. Or achieve both.

  An improved intake manifold for a gas fueled internal combustion engine is in fluid communication with a first distribution chamber in fluid communication with the intake air of the internal combustion engine and at least one intake passage for each combustion chamber of the internal combustion engine. A second distribution chamber, and a diffuser for fluidly connecting the first distribution chamber and the second distribution chamber. The first distribution chamber may have a central inlet and the outer contour of the first distribution chamber may be tapered toward the diffuser on both sides of the inlet. The diffuser comprises a slot having a reduced flow range compared to the first distribution chamber. In a preferred embodiment of the invention, the second distribution chamber is in fluid communication with two intake channels for each combustion chamber.

  Intake ports and valve seat devices for gas fueled internal combustion engines have a difference between the port angle and the valve seat angle that is in the range of -5 ° to + 5 °. The valve seat angle is in the range of 25 ° to 35 °, and in a preferred example, the valve seat angle is substantially 30 °. When the valve member associated with the valve seat is in the open position, the flow in the intake port is forced substantially toward the upper surface of the valve member.

  The improved gaseous fuel internal combustion engine includes a cylinder head, an engine block having a cylinder bore, and a piston associated with the cylinder bore. A combustion chamber is defined by the piston, the cylinder bore, and the cylinder head. A first set of six bolts are arranged to surround the cylinder bore to hold the cylinder head relative to the engine block, preferably in a hexagonal pattern. The first intake passage and the second intake passage extend along one side of the bolt from the intake manifold toward the combustion chamber. In a preferred embodiment of the invention, another cylinder bore and a second set of six bolts disposed around the other cylinder bore are provided. Also, the pair of bolts are made common to the first set and the second set of bolts.

FIG. 1 is a plan view showing a partial cross section of an internal combustion engine having a gaseous fuel combustion apparatus according to a first embodiment of the present invention. FIG. 2 is a plan view schematically showing the internal combustion engine of FIG. FIG. 3 is a perspective view showing an intake manifold, an exhaust manifold, a plurality of cylinders, and intake and exhaust passages of the internal combustion engine of FIG. FIG. 4 is a front view showing the intake manifold of FIG. FIG. 5 is a perspective view showing a gaseous fuel diverter in which one injection valve is fluidly connected to two intake flow paths. 6 is a cross-sectional view partially showing the shunt of FIG. FIG. 7 is a cross-sectional view taken along line 7-7 in FIG. FIG. 8 is a cross-sectional view taken along line 8-8 in FIG. FIG. 9 is an exploded view of FIG. 8 showing one intake valve fully opened. FIG. 10 is a cross-sectional view showing a cross-section across the cross-section of the intake port of FIG. 9 facing the combustion chamber. FIG. 11 shows a Volvo D8K 350 type diesel compression ignition engine having a compression ratio of 17.5 to 1 and displacement of 7.7 liters, and a compression ratio of 12 to 1 and displacement of 7.7 liters. 1 is a graph showing torque curves for one internal combustion engine and a CWI ISL-G type spark ignition natural gas engine having an engine displacement of 8.9 liters. FIG. FIG. 12 is a diagram showing an internal combustion engine having a gaseous fuel combustion apparatus according to a second embodiment of the present invention.

  Referring to the drawings, and initially referring to FIGS. 1 and 2, these figures show an engine 10 having a gaseous fuel combustion device 15 according to a first embodiment of the present invention. The intake manifold 100 has a first distribution chamber 110 and a second distribution chamber 120, also known as a plenum chamber, which are fluidly connected to each other via a diffuser 130. The first distribution chamber 110 is in fluid communication with the throttle 140 to receive air charge from the intake air of the engine 10, and when the engine 10 employs exhaust gas recirculation (EGR), The EGR valve 150 introduces exhaust gas into the air intake flow. A pair of intake passages 20 extend along both sides of one cylinder head bolt 5, and these intake passages 20 pass through the second distribution chamber 120 to the respective intake valves 40 and fluids for the respective cylinders 90. It is connected. Each intake channel 20 has an intake runner 22 connected to a second distribution chamber 120 and an intake port 24 in a cylinder head 240 (best shown in FIG. 8). Although the illustrated embodiment shows six cylinders, other embodiments may have more than one cylinder. As is typical for medium duty engines or large engines, cylinder head bolts 5 are arranged in a hexagonal pattern surrounding each cylinder 90 and two of these bolts are shared between adjacent cylinders. . The exhaust passage 30 extends from the exhaust valve 50, and in the illustrated example, the exhaust passage 30 is connected so as to be an integrally connected exhaust passage reaching the exhaust manifold 160, but departs from the spirit disclosed herein. Other configurations are possible. Each exhaust passage has an exhaust runner 32 and an exhaust port 34 within the cylinder head 240 (best shown in FIG. 8).

  The intake manifold 100 improves the equalization of the charge distribution of air (and EGR) to each cylinder 90 by creating a flow from the first distribution chamber 110 through the diffuser 130 into the second distribution chamber 120. It is designed to have the following characteristics. The outer contour 115 of the first distribution chamber 110 extends toward the diffuser 130 on both sides of the centrally located inlet 105 and extends along the first distribution chamber 110 before entering the second distribution chamber 120. The charge pressure balance is improved. The diffuser 130 is in the form of a slot extending along the first distribution chamber 110 and the second distribution chamber 120. Since the flow range is reduced across the diffuser 130, the charge flow is limited, causing flow impingement on the walls of the first distribution chamber 110 and turbulence in the first distribution chamber 110. Increases flow and overall pressure and achieves a pressure balance along the first distribution chamber 110. The turbulence obtained in the first distribution chamber 110 improves the air-EGR mixing.

  Referring now to FIGS. 5 and 6, the diverter 80 fluidly couples each gaseous fuel injector (injector) 170 (shown in FIGS. 5-8) to each intake flow path 20 via a conduit 85. Each gaseous fuel injector simultaneously introduces gaseous fuel into the paired intake channels for each cylinder 90. The shunt includes a body 82 having a bore into which a gaseous fuel injector 170 is inserted. This bore receives from the gaseous fuel injector 170 and also acts as a plenum and accumulator for the gaseous fuel injected into each intake passage 20 via the conduit 85 by colliding with the end of the bore to increase the pressure. . In the preferred embodiment, the conduit 85 is substantially perpendicular to the longitudinal axis of the bore of the body 82. In the illustrated embodiment, the shunt 80 is an integral component, but in other embodiments, the shunt 80 can be a prefabricated component, in these embodiments, such as a seal. There can be additional components. The gaseous fuel injector 170 is fluidly connected to a gaseous fuel source (not shown) and introduces gaseous fuel into the flow divider 80 when commanded, and the gaseous fuel-air mixture is cylinderd via the associated intake valve 40. To flow into 90. In the present disclosure, reference is made to a gaseous fuel-air mixture, which means that it will also refer to a gaseous fuel-air-EGR mixture depending on engine operating conditions and requirements. Is. The gaseous fuel source supplies gaseous fuel at a pressure suitable for port injection. In the preferred embodiment, the gaseous fuel source accumulates gaseous fuel as compressed natural gas and uses a pressure regulator to reduce the accumulated pressure to a predetermined port injection pressure. A gaseous fuel in the context of the present invention is any fuel that is in the gaseous state at a standard temperature and pressure of 20 ° C. and 1 atmosphere (atm). A typical gaseous fuel is natural gas.

  Each cylinder 90 has a mechanism for igniting a gaseous fuel-air mixture therein. In the embodiment shown, this mechanism is formed by an ignition device 60. In the preferred embodiment, the preferred igniter is a spark plug (as shown in FIG. 7). Referring back to FIG. 1, in the illustrated embodiment, the booster 180 for compressing the intake air is a turbocharger having a turbine 182 and a compressor 184. The outlet of the compressor 184 is fluidly connected to the air intake of the engine 10. In other embodiments, the booster 180 can be a sequential turbocharger, a supercharger, or a combination of these turbochargers and superchargers. The exhaust manifold 160 has an EGR port 155 that is in fluid communication with the EGR valve 150. In other embodiments, the exhaust manifold and EGR device may be similar to that disclosed in co-owned US Provisional Application No. 61 / 870,203. This US provisional application is incorporated by reference.

  8 and 9, each cylinder 90 includes a respective combustion chamber 200 defined by a respective cylinder bore 210, a respective piston 230, and a cylinder head 240 in the engine block 220. Referring to FIG. 9, each intake valve 40 has a respective valve member 42 and a respective valve seat 26 extending annularly surrounding the opening of the intake port 24. As a result, the intake charge flows into the combustion chamber 200. The annular surface 44 on the back surface of each valve member 42 engages with the valve seat 26 to fluidly seal the combustion chamber 200 from the respective intake passage 20 when the respective intake valve 40 is closed. . The valve seat angle α is defined as the angle between the port opening plane 25 and the valve seat. In a preferred embodiment, the valve seat angle α is in the range of 20 ° to 35 °, preferably 30 °. This range of values for the valve seat angle increases durability by increasing tumble motion and reducing wear of the valve seat, as will be described in more detail below. In order to facilitate tumble movement with respect to the combustion chamber 200, the cylinder head 240 has a pent roof 280 that covers the cylinder bore 210, and the piston bowl 270 is generally concave. The slope of the pent roof 280 is similar to the curvature of the piston bowl 270 so that the combustion chamber 230 is approximately in the shape of the object when the piston 230 is at top dead center (TDC).

  The engine 10 is a medium duty engine. In the present disclosure, the diameter of the cylinder bore 210 is defined to be in the range of 90 mm to 120 mm for a medium duty engine. In other embodiments, the diameter of the cylinder bore 210 is greater than 120 mm for larger engines such as, for example, heavy duty engines and engines used in locomotives, mining haul and marine applications. Can be bigger. In the preferred embodiment, the ratio between the diameter of the cylinder bore 210 and the stroke length of the piston 230 (bore to stroke ratio) is in the range of 0.75 to 0.95 so that output is not sacrificed for efficiency. Increases density surprisingly. In practice, efficiency was increased by reducing the heat transfer from the combustion gases to the cylinder bore 210, thereby increasing the energy transfer to the crankshaft of the engine 10. The volume (stroke volume) swept by each piston 230 in each cylinder bore 210 is in the range of 0.8 liter to 2.5 liter. Unlike a light duty engine that uses tumble motion, the maximum engine speed of the engine 10 is 2700 revolutions per minute (rpm) in all operating modes.

  For each cylinder 90, the pair of intake flow passages 20, the respective intake valves 40, and the combustion chamber 200 cooperate to establish a tumble motion of the air-fuel mixture in the combustion chamber. In the preferred embodiment, the average tumble ratio is at least 2. In other preferred embodiments, the average tumble ratio is in the range of 2-5. When the valve member 42 is lifted to the maximum as shown in FIG. 9, the flow (air, EGR, gaseous fuel) along the port floor 290 is pressed in a large amount toward the upper surface 46 of the valve member 42, In this case, a counterclockwise tumble motion is generated in the combustion chamber 200. Any flow below the lower surface 48 of the valve member 42 causes a clockwise tumble motion in the combustion chamber 200 that acts counter to the counterclockwise tumble motion formed by the flow on the upper surface 46. To reduce clockwise tumble movement in the combustion chamber 200, the port floor 290 of the intake port 24 is intentionally spaced from the lower surface 48 of the valve member 42 when the valve is in the closed (seat) position, When in the open position, the flow passes almost over the upper surface 46 of the valve member 42. The port angle θ of the intake port 24, defined as the angle between the port floor 290 and the cylinder cross section 35, reduces non-uniformities and abrupt changes in the flow direction when air enters the combustion chamber 200, and preferably Is equal to a valve seat angle α within a predetermined tolerance to minimize and improve intake port characteristics. In the preferred embodiment, the difference between the port angle θ and the valve seat angle α is less than +/− 5 °. Changes in the direction of flow give rise to a pressure drop that acts as a load block in the direction of flow. Referring to FIG. 10, FIG. 10 shows a cross section of the intake port in which the inside of the combustion chamber 200 can be seen with the valve member 42 completely opened as a cross section along the cross section of the intake port 24. The figure is shown. This figure shows the strong pressing of the flow on the upper surface 46 of the valve member 42 for generating a tumble motion in the combustion chamber 200. The cross-sectional shape of the valve port 24 is generally a square with rounded corners. The port angle θ, the valve seat angle α, and the shape of the intake port 24 cooperate with each other to enhance the tumble motion of the gaseous fuel-air mixture in the combustion chamber 200.

  The turbulent kinetic energy of the gaseous fuel-air mixture is determined by the fact that the tumble motion obtained in the combustion chamber 200 is compressed compared to the swirl air motion combustion chamber and the stationary combustion chamber, and breaks within the turbulent kinetic energy Increase to improve down. The turbulent flame velocity of this mixture increases as well as the local laminar flame front in the turbulent mixture. As the flame speed increases, the knock limit increases, and therefore efficiency can be improved by employing a high compression ratio. The range of the compression ratio is preferably 11: 1 (11: 1) to 15: 1 (15: 1). By operating the engine with EGR cooled at a high EGR rate, the risk of knocking is reduced, and the specific heat ratio of the working gas is increased to improve the Otto efficiency. By making the compression ratio greater than about 15: 1, the heat loss during compression exhibits a diminishing returns state that is greater than the heat that can be recovered by expansion. A compression ignition engine employs a compression ratio greater than 15 to improve cold start performance, but this is not necessary for a spark ignition engine. Reducing the compression ratio compared to a compression ignition engine can reduce the dimensions of the piston and bearing, thus reducing friction and improving efficiency.

  The engine of the illustrated embodiment of the present invention was operated in a test cell fueled with natural gas at a 12: 1 compression ratio and torque data was recorded for a range of engine speeds. The graph diagram of FIG. 11 has a Volvo D8K 350 diesel compression ignition engine with a similar displacement (7.7 liters) but a compression ratio of 17.5 to 1 and a displacement of 8.9 liters. Recorded data comparing torque curves for a CWI ISL-G spark ignition natural gas engine. Surprisingly, it has been found that the engine of the embodiment of the present invention is superior in efficiency to other spark-ignition gas engines (ISL-G) with a considerably larger displacement. Normally, a diesel engine with a high compression ratio would be expected to provide high performance and efficiency. As shown by the curve in FIG. 11, the torque for the engine of the illustrated embodiment of the present invention was superior over the range of engine speeds examined than the diesel engine with a higher compression ratio. The Volvo D8K 350 diesel compression ignition engine (hereinafter referred to as a diesel engine) operates in a lean combustion mode in which the engine 10 operates at or near a stoichiometric air-fuel ratio. The operation of this diesel engine is different from the engine 10 which needs to operate with excess air to avoid exhaust smoke. In this case, the turbocharger of this diesel engine is boosted to provide the required excess air for the same power and torque as the stoichiometric engine 10 which supplies (almost) just air for power and torque. It needs to be significantly larger than the device 180. As a result, the booster 180 employs a turbocharger that is significantly smaller than a diesel engine, so that it can quickly spool up to produce the small boost required by the engine 10 (compared to a diesel engine). Can be. The rotational speed of the engine 10 is kept below 2700 rpm in all operating modes. Operating at lower engine speeds improves fuel economy by reducing the energy lost due to friction. Booster 180 compensates for low engine speed by increasing the pressure in intake manifold 100 and thereby increasing the amount of oxygen available for combustion in combustion chamber 200 for each ignition event.

  Referring now to FIG. 12, this figure shows an engine 11 having a gaseous fuel combustion device 21 according to a second embodiment similar to the first embodiment of the present invention, this second embodiment. Then, the same reference numerals as those in the first embodiment are attached to the same parts as those in the first embodiment, and the detailed description thereof is omitted. The direct injector 65 introduces gaseous fuel directly into the combustion chamber 200 so that the flow divider 80 is not necessary. These direct injectors 65 can be formed in the center of the respective cylinder 90 or can be shifted in the direction of the intake valve 40. Alternatively, these direct injectors 65 can be replaced by injectors (not shown) formed in the wall of the cylinder bore 210.

  According to the technology disclosed herein, a gas fueled internal combustion engine can perform better than a compression ignition diesel engine of similar displacement. This makes the gas-fueled internal combustion engine smaller with respect to displacement compared to conventional internal combustion engines without sacrificing power and torque. In some situations, the number of cylinders can be reduced, thereby significantly reducing the size of the engine and increasing fuel economy.

  Although specific elements, embodiments and fields of application of the present invention have been disclosed and described above, the present invention is not limited to these. The reason is that modifications can be made by those skilled in the art without departing from the scope of the present disclosure, particularly in view of the above-described technique.

Claims (40)

A combustion chamber defined by a cylinder bore, a cylinder head, and a piston that reciprocates within the cylinder bore, wherein the cylinder bore has a diameter of at least 90 mm, and a ratio between the diameter and the stroke length of the piston is increased. A combustion chamber of 0.95,
At least one intake channel for supplying charge to the combustion chamber;
A combustion apparatus for a gas fuel type internal combustion engine, comprising at least one intake valve formed in the cylinder head and causing a significant tumble flow motion in the combustion chamber in cooperation with the intake flow path.   The combustion apparatus of claim 1, wherein the ratio is at least 0.75.   The combustion apparatus according to claim 1, wherein the diameter is 120 mm or less.   The combustion apparatus according to claim 1, further comprising an injection valve for introducing gaseous fuel upstream from the at least one intake valve.   The combustion device of claim 1, further comprising an ignition device disposed within the combustion chamber to facilitate ignition of gaseous fuel and the charge.   The combustion apparatus according to claim 1, wherein the ignition apparatus is an ignition plug.   The combustion apparatus according to claim 1, wherein the flow motion of the tumble has an average tumble ratio in the range of 2-5.   The combustion apparatus according to claim 1, wherein the maximum engine speed of the internal combustion engine is 2700 rpm.   The combustion apparatus according to claim 1, wherein the cylinder bore and the piston define a stroke volume within a range of 0.8 liters to 2.5 liters.   The combustion apparatus according to claim 1, wherein each intake valve has a valve member and a valve seat, and the valve seat has a valve seat angle in a range of 25 ° to 35 °.   The combustion apparatus according to claim 1, wherein the valve seat angle is substantially 30 °.   The combustion apparatus according to claim 1, wherein a difference between the valve seat angle and the port angle is in a range of -5 ° to 5 °.   The combustion apparatus according to claim 1, wherein the compression ratio of the internal combustion engine is at least 11 to 1.   14. A combustion apparatus according to claim 13, wherein the compression ratio is at most 15 to 1. The combustion apparatus of claim 1, wherein the combustion apparatus further comprises an intake manifold, wherein the intake manifold is
A first distribution chamber in fluid communication with the intake air of the internal combustion engine;
A second distribution chamber in fluid communication with the at least one intake flow path;
A combustion apparatus having a diffuser for fluidly connecting the first distribution chamber and the second distribution chamber. The combustion apparatus according to claim 1, wherein
The at least one intake passage is a first intake passage and a second intake passage, and the at least one intake valve is a first intake valve and a second intake valve;
The combustion device is further a shunt that is in fluid communication with the injection valve that receives gaseous fuel, the first intake flow path and the second intake flow path for discharging the received gaseous fuel from the injection valve. And a combustor comprising the shunt in fluid communication with the combustor.   17. The combustion apparatus of claim 16, wherein the shunt has a body having a bore and a pair of conduits, the bores in fluid communication with the injection valve, and each conduit in fluid communication with the bores. And a combustion device in fluid communication with each one of the first intake flow path and the second intake flow path. The combustion apparatus according to claim 1, wherein
The internal combustion engine has an engine block and an intake manifold, and the at least one intake passage is a first intake passage and a second intake passage;
The combustion device further includes six bolts disposed around the cylinder bore, and the bolts hold the cylinder head with respect to the engine block by the bolts. The second intake channel extends from the intake manifold toward the combustion chamber along each side of one of the bolts.   2. The combustion apparatus according to claim 1, further comprising an EGR valve that selectively supplies exhaust gas to the intake manifold.   The combustion apparatus according to claim 19, wherein the exhaust gas is cooled.   2. The combustion apparatus according to claim 1, further comprising a throttle valve that variably supplies air to the intake manifold.   The combustion apparatus of claim 21, wherein the throttle valve is configured to receive a command to maintain a stoichiometric gaseous fuel-air mixture within predetermined tolerances. A gas fuel diverter for diverting a flow of gaseous fuel from a fuel injection valve, the gaseous fuel diverter comprising:
A body portion having a bore in fluid communication with the fuel injector;
A gaseous fuel diverter having a first conduit and a second conduit in fluid communication with the bore;
A gaseous fuel diverter adapted to divert the flow of gaseous fuel into a first flow and a second flow in the first conduit and the second conduit, respectively.   24. The gaseous fuel shunt of claim 23, wherein the fuel injection valve is part of a fuel injector and the bore is configured to accommodate a nozzle of the fuel injector.   24. The gaseous fuel diverter of claim 23, wherein the first conduit and the second conduit are substantially orthogonal to the longitudinal axis of the bore.   24. The gaseous fuel diverter according to claim 23, wherein the main body and the first and second conduits are integrated components. An intake manifold for a gas fuel type internal combustion engine,
A first distribution chamber in fluid communication with the intake air of the internal combustion engine;
A second distribution chamber in fluid communication with at least one intake passage for each combustion chamber of the internal combustion engine;
An intake manifold having a diffuser fluidly connecting the first distribution chamber and the second distribution chamber.   29. The intake manifold according to claim 28, wherein the first distribution chamber has a centrally located inlet, and the outer contour of the first distribution chamber faces the diffuser on both sides of the inlet. Tapered intake manifold.   28. The intake manifold of claim 27, wherein the diffuser comprises a slot having a reduced flow range as compared to the first distribution chamber.   28. The intake manifold of claim 27, wherein the second distribution chamber is in fluid communication with two intake channels for each combustion chamber.   28. The intake manifold of claim 27, wherein the internal combustion engine has a cylinder bore diameter of at least 90 mm and a bore to stroke ratio of at most 0.95.   An intake port and valve seat device for a gas-fueled internal combustion engine, wherein the difference between the port angle and the valve seat angle is in the range of -5 ° to + 5 °.   33. The intake port and valve seat device according to claim 32, wherein the valve seat angle is in a range of 25 [deg.] To 35 [deg.].   34. The intake port and valve seat device according to claim 33, wherein the valve seat angle is substantially 30 degrees.   33. The intake port and valve seat device according to claim 32, wherein when the valve member associated with the valve seat is in the open position, the flow in the intake port is substantially pressed toward the upper surface of the valve member. An intake port and a valve seat device adapted to be used.   33. The intake port and valve seat apparatus according to claim 32, wherein the internal combustion engine has a cylinder bore diameter of at least 90 mm and a bore to stroke ratio of at most 0.95. A cylinder head;
An engine block having a cylinder bore;
A piston associated with the cylinder bore, the piston defining the combustion chamber with the piston, the cylinder bore, and the cylinder head;
A first set of six bolts disposed around the cylinder bore and holding the cylinder head against the engine block;
A gas-fueled internal combustion engine comprising a first intake passage and a second intake passage that extend along each side of one of the bolts from an intake manifold toward the combustion chamber.   38. The gaseous fuel internal combustion engine of claim 37, wherein the six bolts of the first set are arranged in a hexagonal pattern.   38. The gaseous fuel internal combustion engine of claim 37, further comprising another cylinder bore and a second set of six bolts disposed around the other cylinder bore, A gas-fueled internal combustion engine in which a pair of bolts is common to the first set and the second set of bolts.   38. A gas fueled internal combustion engine according to claim 37, wherein the gas fueled internal combustion engine has a cylinder bore diameter of at least 90 mm and a bore to stroke ratio of at most 0.95. .

Images